Piston Pump Having A Force Sensor And A Method For Controlling Said Pump

ABSTRACT

The present invention relates to a piston pump for generating a delivery flow, which is substantially free of pulsation, in particular to a dual piston pump, and to a method for controlling such a piston pump delivering a pumped medium from a low-pressure area into a high-pressure area, wherein measuring sensors detecting mechanical forces or moments transmitted by the structure of the pump or its associated drive unit are used instead of the pressure or flow sensors usually employed for this purpose.

FIELD OF THE INVENTION

The present invention relates to a piston pump for generating a deliveryflow which is substantially free of pulsation, in particular to adual-piston pump, and to a method for controlling such a piston pumpdelivering a pumped medium from a low-pressure area into a high-pressurearea.

BACKGROUND OF THE INVENTION

A piston pump according to the present invention comprises at least twopiston/cylinder units for delivering the pumped medium from alow-pressure area into a high-pressure area, a cam drive for driving atleast one piston/cylinder unit, a control unit for controlling therotational speed of the cam drive, and a sensor for measuring an actualvalue of a control parameter, by which actual value the extent ofpulsation of the delivery flow generated on the high-pressure side canbe derived.

Pumps of this kind are employed, for example, in liquid chromatography,especially in high-pressure liquid chromatography (HPLC) and inultrahigh-pressure liquid chromatography (UHPLC), for the delivery ofthe mobile phase (the eluant), for example, in form of a pure solvent ora low pressure side gradient solvent mixture through the stationaryphase (package) in a separation column of a pertinent analysis system.The term “low pressure side gradient” will be familiar to each skilledperson.

SUMMARY

Correct analysis results with regard to quality (retention time) andquantity (peak area) require a continuously constant mass flow of themobile phase. Ideally, true mass flow should be ensured, i.e. a deliveryflow, which is constant over a certain unit of time with respect tovolume at atmospheric pressure. A true mass flow should be ensured atleast for the duration of a series of analysis runs associated with eachother. A constant mass flow is of critical significance, especially inHPLC and UHPLC, since the delivery pressures of said pumps may reach 100MPa and more when using said analysis methods.

Under such operating conditions, the pumped medium no longer behaves asan ideal, i.e. incompressible liquid. This results in the fact that dueto the specific compressibility of the pumped medium a rising deliverypressure increasingly exacerbates the generation of true mass flow ofthe pumped medium, which is substantially free of pulsation.

In order to reduce pulsations in a high-pressure delivery flow generatedby a piston pump, for example, pumps having a plurality of, inparticular two, pistons are employed operating in accordance with theparallel or preferably the serial delivery principle. In other words,the pistons deliver into a common high-pressure area, wherein theirstroke movements are offset in time with respect to each other such thatthe individual delivery flows are theoretically superimposed over eachother to form a composite constant flow, which is free of pulsation.

With low delivery pressures a pulsation of the total delivery flow canthus be completely or nearly avoided, with high and highest deliverypressures, however, a residual pulsation arises. This residual pulsationis caused by the specific compressibility of the pumped medium, whichhas an increasing effect as the delivery pressure rises and its extentdepends on the changing physical properties of the pumped medium. Saidresidual pulsation may be reduced or suppressed by commencing eachdelivery cycle with a precompression stroke, which ensures that themedium within the cylinder internal space is compressed to match therespective system pressure prevailing on the high-pressure side prior tothe onset of the actual delivery in order to avoid or at least minimizea short-term backflow of pumped medium and/or an interruption of thecontinuous delivery upon hydraulic transition from one piston to theother.

Feedback controls, known from pertinent pumps, for (continuously)compensating the influence of the specific compressibility of the pumpedmedium on the momentary pumping efficiency are either based onmonitoring the delivery pressure or determining the flow rate (e.g. bysensors operating according to the caloric measuring principle) at theinlet and/or outlet of a pump as well as the qualitative or quantitativeprocessing of the respective measuring signals.

U.S. Pat. No. 4,359,312 discloses a dual-piston pump having a cam driveoperating in accordance with the parallel or serial delivery principleof the kind described above as well as a method for controlling saidpump based on a subtractive approach. In order to be able to counteracta drop in the pumping efficiency at increased delivery pressures, thedrive cam(s) causing the actual pumping action comprises an elevationprofile section for generating a pre-compression stroke, with a suctionstroke correspondingly shortened in relation to the angular range andaccelerated due to construction. Initially, the length of thepre-compression stroke is chosen such that it compensates for theinfluence of the specific compressibility of the pumped medium on thepumping efficiency at a specified maximum specific compressibility and aspecified maximum allowable delivery pressure the pump is driven so asto fully compensate for these maximum conditions. For this purpose thedevice comprises a pressure sensor by means of which the system pressureon the high-pressure side is continuously monitored.

The measured values monitored by the pressure sensor are fed into thedrive control system and the rotational speed of the drive motor ismodulated in the pre-compression stroke range such that, with aniterative and subtractive approach, the pump is iteratively backed offfrom compensation of the maximum conditions so as to compensate for theactual conditions of the pumped medium at the actual operating pressure.In parallel therewith, an increased delivery with respect to volume atatmospheric pressure is obtained due to the pre-compression stroke usedrespectively in the feedback process. This surplus delivery is correctedby means of readjustment by superimposing a secondary correction factorto the cam rotational speed.

In U.S. Pat. No. 4,681,513 another pump and another method aredisclosed. Herein, each of the two pistons of a dual-piston pumpoperating in accordance with the serial delivery principle has aseparate cam drive. Each of the two drive cams, in sections, hasdifferent elevation profiles characterizing different stroke lengths,which are graded according to certain maximum delivery rates of thepump, with their slopes exhibiting a constant rise. During each pumpingcycle, within the associated angular range of the cams, both, thesuction stroke and the pumping stroke as well as a precedingpre-compression stroke are caused by a reciprocating movement of thedrive cams and the two delivery flows are combined such that a compositeconstant delivery is achieved.

A crucial inherent disadvantage of all feedback controls forcompensation of compressibility (pressure feedback), which are based ona qualitative and/or quantitative utilization of the measurement of thedelivery pressure, is the need for an iterative adjustment (PID control)within a series of cam revolutions/pumping cycles and a practicallylimited robustness of the control upon occurrence of pressure artefacts,since only an indirect and no direct monitoring of the process ofdisplacement of the pumped medium is possible by the pressure measuringsignal. Disturbance variables are, for example, pressure pulses,occurring when a sample is fed into the pertinent analysis system, andin particular the drift of the pressure measuring signal during gradientoperation, wherein back-pressure changes arise or may arise at theseparation column due to varying physical properties (viscosity;temperature) and/or when a clogging of the column occurs or may occur.

Apart from the pressure feedback method for compensating the influenceof the specific compressibility of the actually pumped medium on thepumping efficiency, a flow feedback method is known, which is based uponmonitoring the delivery characteristics of the pump by means of flowsensors arranged on the high-pressure side (and optionally also on thelow-pressure side) within the liquid duct path and wherein the feedbackmethod relies on measuring signals of said sensors.

Specific technical problems arise with the pressure measuring sensors aswell as with the flow sensors from the fact that they must be suited toa very wide measuring range of 0-100 MPa for example and thus, must beable to withstand the high pressure and be peripherally sealable withacceptable effort. At the same time, they must not affect optimalconfiguration of the geometry of the liquid paths in terms of pressuredrop and fluid dynamics and, since they are in direct contact with themedium to be delivered, they must either be chemically resistant to allmedia to be delivered or be protected from said media by a separationmembrane or the like. In the case of flow sensors it is of particulardisadvantage that the measuring signal is dependent upon the operatingtemperature as well as the specific thermal conductivity/capacity of therespective medium to be delivered and that consequently, amultiparametric calibration with regard to application is necessary,which can be achieved during operation of the pump in accordance withthe low-pressure gradient method at best by a detour via the method of alearning-in run by which a set of reference values is assessed underreal analysis run conditions.

The difficulties described above with regard to design conceptsuperimpose the basic systematic problem of the decrease of the pumpingefficiency by detrimental dead volume within the displacement system(piston/cylinder unit) of the working piston when considering the caseof a serial dual-piston pump. Said volume being that which is notdisplaced from the displacement chamber—calculated (ranging) from theclosing edge of the inlet valve to the closing edge of the outletvalve—during the pumping stroke and which, remaining therein, acts as ahydraulically elastic element, since liquids have a noticeable specificcompressibility. Implementing the concept of all-in/all-out, whichtheoretically presents a solution to the problem, has technicallimitations as to the respective design of the components belonging tothe displacement chamber feasible in practice.

With increasing delivery pressure, the compressibility of the medium tobe delivered causes an increasing drop of the pumping efficiency andassociated with that a delivery flow having a more or less pronouncedresidual pulsation. The pumping efficiency is diminished solely as afunction of the detrimental dead volume, which is technically difficultto minimize, and not as a result of the compression of the (actual)pumping volume during the displacement stroke.

When using the pump while applying the low-pressure gradientmethodology, the detrimental dead volume also affects theproportionating of the individual feed flows caused by theproportionating valves on the suction side of the pump. This is due tothe fact that, during the onset of the suction stroke, said dead volumehas first to expand to the volume at ambient pressure before “fresh”liquid can flow into the displacement area. This causes a reduction inthe filling stroke efficiency a problem not addressed by the pumps ofthe prior art discussed above.

Based on the prior art described above, it is the object of the presentinvention to provide a piston pump as well as a method for controllingsaid pump, wherein, even with varying specific compressibility of themedium to be delivered and changing delivery pressures of up to 100 MPaand possibly more, the full pumping efficiency is maintained and a(residual) pulsation of the delivery flow is reduced or even suppressedcompared to known pumps and methods.

As regards the device, this object is achieved by a piston pumpcomprising at least two piston/cylinder units for delivering the mediumout of a low-pressure area into a high-pressure area, a drive unit fordriving at least one piston/cylinder unit, a control unit forcontrolling the rotational speed of the drive unit and a sensor forcollecting an actual value of a control parameter, by which actual valuethe extent of pulsation of the flow generated on the high-pressure sideis derivable, the piston pump being characterized in that the sensor isadapted for collecting values of mechanical forces and/or torquesexerted and/or transmitted in/at the structure of at least onepiston/cylinder unit and/or of the drive unit.

By this design of the pump according to the invention, the (otherwise)compulsory use of pressure or flow measuring sensors is advantageouslyavoided. Thus, no sensor has any longer to be arranged in the areawetted by the pumped medium, i.e. in the high-pressure area and the zonewetted by that medium. Rather, it can be arranged in almost any locationand is only required to be suitable to detect the mechanical forces andmoments exerted and/or transmitted within the structure of the pumpduring operation of the pump. Preferably, the sensor is arranged suchthat it is not wetted by the medium to be delivered, especially not bythe already pre-compressed or compressed medium.

By structure of the pump basically all mechanical assemblies andcomponents of the pump are meant. In particular, the sensor monitorsforces and moments of the aforementioned kind, which are transmitted byor exerted in/at the structure of the piston/cylinder unit, the driveunit or the components interfacing the drive unit and thepiston/cylinder unit. The drive unit of the pump in the above sensecomprises transmission and drive units. Advantageously, it is a camdrive and comprises associated drive shafts including cam disks as wellas the above transmission and drive units.

The described arrangement of the sensor yields several advantages.

By means of this, in comparison with conventionally structured pumps,the transfer volume is significantly reduced, which is decisive forputting low-pressure gradients through the pump, preferably withoutre-mixing and, consequently, for the suitability thereof in providing ahigh sample throughput (HTP) by the pertinent analysis system. Thetransfer volume is defined as the volume in the total liquid duct pathranging from the closing edge of the proportionating valves forming thegradient to the fitting at the outlet of the pump. The smaller thetransfer volume, the less time is required for resetting the analysissystem to initial conditions for the next analysis run.

If no sensor has to be arranged in the zone wetted by the pumped medium,there is no need for integrating sensors, which are usuallytwo-dimensional rather than desirably tubular-shaped and which also haveto be chemically inert depending on their use, which integration in afluid-tight manner at high pressure is technically complicated andcostly. This has the favourable side effect that also the geometries ofpassages can be configured in a simple manner such that only directflow-through liquid areas are created.

Moreover, the concept of monitoring by means of force measuring withinthe mechanical zone not wetted by the pumped medium inherently offersthe advantage of a stroke synchronous monitoring of the liquiddisplacement action, which can be accomplished (even) resolved within apumping cycle in a functionally robust manner. In contrast toconventional methods which are based on measuring the delivery pressureand the flow rate, delivery pressure artefacts described earlier, whichmight arise due to a change of the viscosity of the pumped medium or achange of temperature thereof, pressure surges when a sample is fed intothe following analysis system, drift of the pressure measuring signalduring low pressure gradient operation and clogging of the followingcolumn, do not constitute interference factors having a direct effectsince not the absolute value of a change but the rate of a change isprocessed:

Since, with said type of sensor, the measurement of mechanical forcesand/or moments exerted/transmitted in/at the mechanical assemblies andcomponents of the piston pump takes the place of a measurement providingonly an indirectly measured value (e.g. a direct measurement of thedelivery pressure), the accordingly monitored values, apart fromcontrolling the pump, may also be evaluated for further functions, interalia for checking the (ball) valves at the liquid displacement unit(s)for proper functioning and/or for checking the piston seals for theextent of wear present.

The sensor is preferably one which is adapted to monitor tensile,compression, shear and/or torsional stresses in and/or at the pumpstructure or components embedded or integrated into it including thedrive unit. In principle, this may be a sensor monitoring mechanicalforces and/or moments, which especially comprises a strain gauge, apiezo-electric element, an acoustic resonator or an optical measuringdevice.

For example, a torque or torsion sensor may be used, which is engaged inor at the shaft of the drive unit or cam drive and monitors thetorque/torsional moment exerted or transmitted there. Another way is tomonitor the torque transmitted at the drive unit of the cam drive bymeans of the sensor. It is of particular advantage that the sensor isnot arranged in the area of the pump wetted by the pumped medium on thehigh-pressure side. Advantageously, even a wetting of the sensor by themedium on the low-pressure side is also avoided.

The piston pump according to the present invention is preferably aparallel or serial dual-piston pump. In accordance with a particularlyadvantageous embodiment as a serial dual-piston pump, thepiston/cylinder units are formed as a working-piston unit, especiallyhaving inlet and outlet valves, and as a storage-piston unit, which areconnected to each other in such a way that the working-piston unit drawsin the medium to be delivered from the low-pressure area of the pump andprovides said medium to one portion of the high-pressure area suppliedby the pump on the outlet side and to the other portion on the inletside of the storage-piston unit in order to achieve a constant deliveryflow by means of stroke movements adjusted to each other and such thatthe storage-piston unit provides the full delivery rate during thesubsequent suction stroke of the working-piston unit.

The piston pump preferably comprises one drive unit each for eachindividual piston/cylinder unit. Moreover, each drive unit can be drivenby a separate motor. This results in a high degree of technical freedomwith regard to controlling and monitoring the delivery volumes of theindividual piston/cylinder units, since both drives may be adjusted toeach other within wide limits by modulating the rotational speed of boththe motors.

As regards the method, the object of the present invention is achievedby a method for controlling a piston pump for delivering a pumped mediumout of a low-pressure area into a high pressure area, preferably adual-piston pump, which piston pump comprises at least twopiston/cylinder units, at least one drive unit for driving at least oneof the piston/cylinder units and a sensor for monitoring mechanicalforces and/or torques transmitted and/or exerted in/at the structure ofat least one piston/cylinder unit and/or the drive unit, which methodcomprises the following steps:

i) driving at least one piston/cylinder unit by a drive unit at a firstspeed (n), thereby monitoring mechanical forces and/or torquesexerted/transmitted in/at the structure of said piston/cylinder unit(s)and/or said drive unit by the sensor,

ii) monitoring the moment of the actual onset of delivery of the pumpedmedium out of the piston/cylinder unit(s) into the high-pressure area byusing the values of the mechanical forces and/or torques monitored bysaid sensor,

iii) monitoring the rate of compression at the moment of the actualonset of delivery, particularly the presence of over- orunder-compression (so called compressibility compensation),

v) modulating or adjusting the drive unit rotational speed to a secondspeed (n+1), in such a way that a varying compression rate due tovarying system back pressure and/or varying compressibility of thepumped medium arising on the high-pressure side is compensated and insuch a way that an essentially pulse-free delivery flow is generated inthe high-pressure area,

v) repetition of method steps i) to iv) for each pumping cycle by usingthe drive unit rotational speed (n+1) modulated in step iv) as the firstspeed (n).

The method according to the present invention is preferably recursivelypursued for each pumping cycle. However, a control cycle can also bebased on the measured data of several delivery strokes. The methodenables the generation of a high-pressure delivery flow over a widedelivery pressure and specific liquid compressibility range, saidhigh-pressure delivery flow being essentially free of pulsation. Inprinciple, one or more piston/cylinder units can be monitored bysensors. This is to be considered when reference is made solely to onepiston/cylinder unit for the sake of convenience throughout the presentspecification.

When starting-up the pump upon the onset of the pumping operation, i.e.shortly after power on, the pump first has to pass through a start-upphase until a steady delivery flow is reached after its initialstart-up. Basically, the drive speed may have an arbitrary value duringthe start-up phase; it is, however, preferably adjusted to the specificcompressibility of the range of pumped media to be expected as well asto the desired rate of delivery, which facilitates the subsequentcontrol of the pump.

After reaching a steady delivery flow or a stable delivery behavior ofthe pump, the actual method according to the present inventioncommences. According to this method, the pump is driven at a first speed(n) at least during one pumping or delivery cycle. Preferably, themechanical forces and/or moments transmitted/exerted in/at the structureof the pump are continuously monitored by the sensor. In particular,forces and/or moments are monitored, which are transmitted and exerted,respectively, in/at the structure of the piston/cylinder units, thedrive unit or by intermediate units. In accordance with a particularembodiment, control programs can be employed for this purpose, which areadapted to the respective rate of delivery chosen.

In the subsequent method step ii), the time and thus the piston positionis determined by using the mechanical forces and/or moments monitored bythe sensor, at which time the actual delivery of the medium into thehigh-pressure area sets on while driven at the first speed (n) or atwhich piston position this happens at least at the piston/cylinderunit(s) monitored by the sensor. The onset of the actual delivery can bemonitored unambiguously and particularly well by using the course of themechanical forces and/or moments monitored by the sensor, for example bymathematically deriving the gained measured values or series of measuredvalues, e.g. with respect to time or place position (piston strokeposition).

While driving the piston/cylinder unit at the first speed (n), systempressure builds up within the pertinent cylinder displacement chamber asthe delivery stroke of the piston increasingly advances. This pressureincreases until it is equal to the system pressure prevailing in theadjacent high-pressure area. At the time at which there is a pressurebalance within the high-pressure area and within the cylinder space ofthe piston/cylinder unit, the medium is not yet delivered from thecylinder space. Only after the balance between the pressure within thecylinder space of the piston/cylinder unit and the system pressure onthe high-pressure side has been exceeded can the actual delivery takeplace.

Considering the mechanical forces and/or moments monitored by thesensor, the onset of the actual delivery becomes recognizable in thatthe course of the monitored force and/or moment measuring values changessignificantly. The utilization of the courses of the force and/or momentmeasuring values monitored by the sensor can be advantageouslyfacilitated by additionally or alternatively assessing and monitoringthe first and/or second stroke position dependant derivations of thesevalues or courses.

Since the kinematic characteristics of the drive unit, in particular ofdrive cam(s) and the (respective) speed (n) of the cam drive(s) areknown, the rate of (pre-) compression of medium previously effectedwithin the piston/cylinder unit(s) can be calculated by using the strokeposition detected by the sensor or the specific drive unit position (camangle), at which the actual delivery has set on. In particular, thepresence of an over or under compression (over- or under compressibilitycompensation) can be determined. In other words, it can be determined,whether the (actual) delivery starts too early or too late with regardto the respective geometry of the drive unit (cam drive) present andwhether a (residual) pulsation of the generated high-pressure flowarises or may arise as a result. In the case of an angular rangeintegrated in the drive unit or cam drive specially for generating apre-compression stroke it can thus be determined, whether, at therespective delivery pressure and/or the respective specificcompressibility of the pumped medium, the compression caused byemploying said cam section at given drive speed over or undercompensates the influence of the specific compressibility of the pumpedmedium on the pumping efficiency, and thus results in a residualpulsation of the delivery flow. This feature is of particularsignificance when operating the pump according to the low pressure sidegradient method (supplying the medium on the suction side with acomposition changing as a whole).

Determination of the (pre-) compression effected prior to the onset ofdelivery is preferably performed in a stroke-related manner. Inparticular, it is checked whether an over or under compensation of thespecific compressibility of the pumped medium is present by using themonitored pulsation characteristics.

In the further process of the method, the drive unit rotational speed(or cam drive speed) is modulated or adjusted to a second speed (n+1).Said speed may be higher or lower than the previous speed (n), which isdependent on the actual required pre-compression determined in theprevious pumping cycle. As a result of this modulation, a previouslydetermined too high or too low rate of pre-compression (over or undercompensation) is compensated for or its extent is at least minimized. Asa consequence, despite of changed system pressure on the high-pressureside and/or a changed specific compressibility of the pumped medium, adelivery flow substantially free of pulsation is generated.

The second speed (n+1), which is adapted to the changed operatingparameters, is used as the first speed (n) in step i) in the furtherprocess of the method and the method steps described above arerecursively pursued. Preferably, this is done already during therespective subsequent piston stroke.

Based on a reference value derived in the described manner, with thedescribed method, it is possible to compensate the influence of changingcompression ratios from one stroke to the subsequent stroke, whichratios arise as a result of the changed specific compressibility of thepumped medium and/or the changed delivery pressure caused by thevariable system pressure on the high-pressure side.

The drive unit rotational speed (cam drive speed) preferably remainsconstant during an initial delivery stroke, more preferably during a(initial) delivery stroke initiating the described control method. Inparticular, the monitoring of the mechanical forces and/or moments orthe course thereof is performed during a delivery stroke at a constantdrive speed.

As regards the method, it is of particular advantage in the methodaccording to the present invention, if the medium to be delivered ispre-compressed in each pumping cycle. The pre-compression strokerequired for that is preferably adapted to the generation of a maximumdelivery pressure and an expected maximum specific compressibility ofthe medium to be delivered. The pre-compression generated by said strokeis sufficient for at least one specified maximum delivery pressure andfor an application-specific expected maximum specific compressibility ofthe fluid to be delivered. Consequently, it is to be expected thatduring operation at the first speed (n) an over-compression will arisein method step i), whereupon the drive unit rotational speed is reducedby modulation during the pre-compression stroke. This is described as aniterative subtractive approach to pump control elsewhere in thisdescription.

According to a particular embodiment of the method according to thepresent invention it is foreseen to derive a correction factor by meansof the stroke position of the delivery piston determined (by thesensor), in which position the actual delivery sets on. By means of thiscorrection factor a running program adapted thereto is generated orchosen from a plurality of stored running programs, which then forms thebasis for the modulation of the drive unit rotational speed, especiallywithin the cam section relevant for the pre-compression stroke.

By means of the described pre-compression and/or modulation of the drivespeed (primary control) which results in a modulation of the pistonspeed, an excessive delivery in relation to the previously chosennominal delivery amount will take place. In particular, an excessivedelivery in relation to a (supplied) volume at ambient pressure willtake place as initially; the pump is driven to compensate for maximumcompressibility and maximum operating pressure. According to a furthersuggestion of the present invention, this excessive delivery iscompensated for by superimposing a secondary correction factor onto themodulated (cam) drive speed, which correction factor provides for acompensation of the excessive delivery (secondary control: generatingtrue mass flow).

In accordance with a particular embodiment of the method according tothe present invention, the values or value patterns detected by thesensors are used to correct the efficiency loss in relation to thesuction stroke, which arises at the beginning of each suction stroke byexpansion of the detrimental dead volume in the displacement area of thepiston/cylinder unit compressed during the previous delivery stroke inaccordance with the respective specific compressibility of the pumpedmedium and in accordance with the respective delivery pressure, i.e. thevolume, which, due to the design, is inevitably not displaced during thedelivery stroke, however, must be compressed prior to the onset of theactual delivery. This is preferably done by a determination of theduration of the decompression phase. The portion of the stroke volumewhich cannot be displaced from the displacement area of thepiston/cylinder unit during each pumping cycle due to expansion of thedetrimental dead volume is monitored. The non-usable portion of thesuction stroke results from the expansion of the detrimental deadvolume. In particular, during low pressure gradient operation, i.e. whenthe composition of the medium to be delivered changes and subsequentlyvarying compressibility and/or viscosity affects the back-pressure. Thisis of particular advantage at very high delivery pressures, since thegradient composition of the medium flowing into the pump during thesubsequent suction strokes can be adjusted in accordance with the extentof expansion of the detrimental dead volume. Preferably, this is done byopening and closing the solenoid valves usually employed for thispurpose in an adaptively controlled manner at the inlet of the pump foreach subsequent suction stroke.

BRIEF DESCRIPTION OF THE DRAWINGS

Further advantages and features of the present invention are apparentfrom the following description of a non-restrictive embodiment withreference to the figures, in which:

FIG. 1 shows a schematic illustration of the pump according to thepresent invention including a control unit in an apparatus for(high-pressure and ultrahigh-pressure) liquid chromatography,

FIG. 2 shows diagrams, in which the normalized strokes of the workingpiston (upper diagram) and of the storage piston (lower diagram), areillustrated as a function of the angle of rotation of the cam shaft,

FIG. 3 shows diagrams, in which the normalized delivery rate of theworking piston (upper diagram) and of the storage piston (lower diagram)are illustrated as a function of the angle of rotation of the cam shaft,

FIG. 4 shows diagrams, in which the normalized forces present at theworking piston (upper diagram) and at the storage piston (lowerdiagram), are illustrated as a function of the angle of rotation of thecam shaft,

FIG. 5 shows a diagram, in which the alteration of the force present atthe working piston is illustrated as a normalized first derivative, as afunction of the angle of rotation of the cam shaft, and

FIG. 6 shows diagrams, in which the open condition of the inlet valve(upper diagram) and of the outlet valve (lower diagram) are illustratedas a function of the angle of rotation of the cam shaft.

FIG. 7 a,b show diagrams, in which for the most preferred embodiment thenormalized stroke of the working piston (FIG. 7 a) and of the storagepiston (FIG. 7 b), are illustrated as a function of the angle ofrotation of the cam shaft,

FIG. 8 a,b show diagrams, in which for the most preferred embodiment thenormalized delivery rate of the working piston (FIG. 8 a) and of thestorage piston (FIG. 8 b), are illustrated as a function of the angle ofrotation of the cam shaft,

FIG. 9 shows a diagram, in which for the most preferred embodiment thenormalized cam rotational velocity is illustrated as a function of theangle of rotation of the cam shaft,

FIG. 10 a.b show diagrams, in which for the most preferred embodimentthe normalized forces present at the working piston (FIG. 10 a) and atthe storage piston (FIG. 10 b), are illustrated as a function of theangle of rotation of the cam shaft,

FIG. 11 shows a diagram, in which for the most preferred embodiment thenormalized first derivation of the force at the working piston isillustrated as a function of the angle of rotation of the cam shaft,

FIG. 12 shows the principal layout of the dual stage gear system,

FIG. 13 shows the gear system including one of the two Z-shaped drivearms, and

FIG. 14 shows a lengthwise cut in flow direction of the most preferredembodiment of the liquid end.

DETAILED DESCRIPTION OF EMBODIMENTS

An exemplary embodiment of the piston pump according to the presentinvention is shown in FIG. 1. The pump 1 designed according to theserial delivery principle comprises a working piston/cylinder unit 2 anda storage piston/cylinder unit 3. The working piston/cylinder unit 2essentially consists of a working piston 4 performing a reciprocatingmovement within the cylinder space of a working cylinder 5. Similarly,the storage piston/cylinder unit 3 consists of a storage piston 6performing a reciprocating movement within the cylinder space of astorage cylinder 7. The working piston 4 and the storage piston 6 areeach driven by a separate drive unit. As is apparent from FIG. 1, bothdrive units are identical, therefore, only the drive unit of the workingpiston 4 is explained in detail in the following. The explanationsequally hold true for the storage piston 6.

The drive unit of the working piston 5 comprises a motor 8. Inprinciple, any kind of motor may be used as motor 8, however, aconventional electric motor, such as a direct-current motor, amagnetostrictive or a piezo-electric drive system is preferred. Themotor 8, in combination with an absolute encoder unit 9, is coupled tothe gear 10 featuring a drive cam. The position of the drive cam withinthe gear 10 is monitored and associated in an exact and correctlypolarized manner in relation to the stroke movement of the workingpiston 4 by means of the absolute encoder unit 9. This may also be done,for example, by means of index disks or rotary encoders in combinationwith specialized control software.

The motor 8 is controlled by a servo unit 11, which in turn is connectedto a computing unit 12. A control program and preferably, in table form,specific control programs for modulating or adapting the rotationalspeed of the motor in accordance with various compression correctionfactors are stored inter alia in the computing unit 12 or are generatedcomputationally.

Between the gear 10 and the pertinent piston (working piston 4 orstorage piston 6) one force/moment sensor 13 each is arranged both in orat the drive of the working piston/cylinder unit 2 and in or at thedrive of the storage piston/cylinder unit 3. The force/moment sensor 13monitors forces and/or moments exerted or transmitted between therespective piston and the transmission unit. The gear unit 10 isconfigured as a cam drive unit. The cams of said unit basically may havearbitrary kinematic profiles, however, an angular range for generating apre-compression stroke initiating the delivery stroke and adapted tomaximum specified operating conditions with regard to delivery pressureand) to medium compressibility is always provided. Said cam section isemployed in accordance with the extent of the system pressure in analiquot manner by maintaining a basic drive speed, which is modulated inthe surplus section not used according to definition such that thestroke movements of the working piston and the storage piston generate acomposite constant delivery flow in accordance with the set deliveryrate.

The computing unit 12 is connected via the servo unit 11 to both thedrive of the working piston/cylinder unit 2 and the drive of thestorage-piston/cylinder unit 3. As indicated by the dashed lines in FIG.1, the computing unit 12 is further connected to the absolute encoderunit(s) 9 as well as to the force/moment sensor(s) 13 and monitorsand/or processes the measuring signals thereof.

The working-piston/cylinder unit 2, on its inlet side, comprises aninlet valve 15 at its inlet 14 acting as an uncontrolled check valve. Aproportionating valve unit 16 is arranged before said inlet valve beingfitted with four special solenoid valves for low pressure side gradientformation. Said unit is, in turn, connected to the computing unit 12 forcontrol purposes and is controlled by said computing unit. Duringoperation of the pump, medium to be delivered can be withdrawn fromseveral respective reservoirs 17 a-d by suitably controlling theproportionating valve unit 16. Gradient formation on the suction side isaccomplished by conducting medium alternately from the reservoirs orsources of medium 17 a-d via the inlet valve 15 into the workingcylinder 5 by a programmed control of the solenoid valves of theproportionating valve unit 16 during the suction stroke of the workingpiston 4. During the delivery stroke of the working piston 4 followingthe suction stroke, the medium drawn into the working cylinder 5 isdelivered via an outlet 18 of the working-piston/cylinder unit 2,through an outlet valve 19 also configured as an uncontrolled checkvalve and via the inlet 21 into the storage piston/cylinder unit 3,which is without valves in the case illustrated.

The delivery flow from the working-piston/cylinder unit 2 to thestorage-piston/cylinder unit 3 flows into the storage cylinder 7 via theinlet 21. The stroke movements of the working piston 4 and the storagepiston 6 are adapted to each other in such a way that the storage piston6 simultaneously performs its suction stroke during the delivery strokeof the working piston 4. Accordingly, the working piston 4, on the onehand, delivers into the storage-piston/cylinder unit 3 and, on the otherhand, through this unit into the high pressure line 20 forming the feedline to the system supplied with the medium. With the present functionalprinciple, no further compression of medium takes place in the storagepiston/cylinder unit 3. The medium is always under system pressuretherein. Compression as well as pre-compression is performed exclusivelyin the working piston/cylinder unit 2. The storage-piston/cylinder unit3 only serves as a storage and delivery reservoir for bridging theinterruption of delivery of the working piston/cylinder unit 2 duringits suction stroke. The storage-piston/cylinder unit 3 may form areservoir having a volume which can be adapted to the respectiveoperating conditions. The volume stored in the storage piston/cylinderunit 3 is delivered from the storage piston/cylinder unit 3 into thehigh-pressure line 20 during the suction stroke of theworking-piston/cylinder unit 2.

In functional continuation of the system, a high-pressure sampleinjection valve 24 is arranged in the high-pressure line 20 or behindthat line in front of a separation column 23. At the outlet of theseparation column 23 a detector 28 is arranged, by means of which thechemical compounds can be detected which are injected with the samplevolume and being eluted by differential retardation from the separationcolumn by the delivery flow according to their differential partitionbetween the separating phases. Medium conducted through the separationcolumn 23 and the detector is received in a waste container 29.

The pump unit depicted in FIG. 1 further shows, e.g. for safety reasons,a pressure sensor 25 monitoring the medium pressure (system pressure)prevailing in the high-pressure line 20, the values of which aretransmitted via the signaling line, illustrated by a dashed line, to thecomputing unit 12 for [back-up] control purposes. It must be noted thatthe additional pressure sensor 25 is optional and not compulsory butrepresents an optional control device.

In FIGS. 2 to 6 exemplary operating diagrams of the piston pumpillustrated in FIG. 1 are shown. In each of the diagrams, the angle ofrotation of the cam shaft is plotted on the abscissa within a range from0° to 360°, wherein specific rotation angle positions to be consideredare highlighted.

Point A marks the angle of rotation of the cam shaft, at which thepre-compression phase of the working piston 4 is completed. Bcharacterizes the angle of rotation, at which delivery of medium intothe high-pressure area is affected solely by the working piston 4. Ccharacterizes the angle of rotation, at which the exclusive delivery ofmedium solely by the working piston 4 terminates. D characterizes theangle of rotation, at which the medium begins to flow from thelow-pressure area into the cylinder space of the working piston/cylinderunit 2 upon onset of the suction phase. Each of the diagrams of FIGS. 1to 6, on its abscissa, shows a full revolution of the cam shaft (from 0°to)360°. After employing the cam section from the angle of rotation 0 topoint E, which corresponds to a complete revolution of the cam shaftover 360°, the entire working cycle commences again at position 0. Atwhich point the delivery stroke of the working piston 4 commences (upperdiagram, FIG. 2).

The stroke of the working piston 4 performed from 0 to A generatespre-compression only. Both the inlet valve 15 and the outlet valve 19remain closed during employing this stroke range (FIG. 6). The storagepiston 6 performs part of its delivery stroke within the same rotatingrange of its drive cam (lower diagram, FIG. 2). As can be seen from FIG.3, delivery of medium into the high-pressure line 20 over this rotationangle range is solely performed by the storage-piston/cylinder unit 3.The delivery rate achieved by the working piston/cylinder unit 2 iscompletely consumed for pre-compression (upper diagram, FIG. 3)—thepiston displacement action does not deliver pumped medium, it compressesit. FIGS. 4 to 5 show the normalized forces present at the workingpiston 4 as well as at the storage piston 6 and the normalizedalteration of the forces present at the working piston 4 (the firstderivative at cam stroke position of the normalized force). The upperdiagram of FIG. 4 shows that the force present at the working piston 4rises in an even and, if necessary, constant manner within the rotationangle range from 0 to A. This may also be gathered from FIG. 5illustrating an alteration of the force present at the working piston 4.

When point A of the angle of rotation of the cam shaft is reached, themedium within the cylinder space of the working-piston/cylinder unit 2is pre-compressed in accordance with the system pressure on thehigh-pressure side. This means that the pre-compression phase iscompleted. The upper diagram of FIG. 6 shows that the outlet valve 19opens at this point of time. The stroke of the working piston 4 advanceswith unchanged speed just as in the pre-compression phase between 0 andA (upper diagram, FIG. 2), whereas the stroke of the storage piston 6 isslightly delayed as compared to the course of the pre-compression phasebetween 0 and A (lower diagram, FIG. 2). FIG. 3 shows that in the rangebetween A and B the normalized displacement action of the storage piston6 slightly drops in comparison to that during the pre-compensation phasebetween 0 and A by the amount of the normalized displacement action ofthe working piston 4, which corresponds to the representation in theupper diagram. The displacement action of the working piston 4 is nolonger consumed for pre-compression in the range between A and B. Thisresults in the fact that both the storage piston 6 and the workingpiston 4 deliver medium into the high-pressure line 20. As shown in theforce diagrams of FIGS. 4 and 5, the normalized force present at theworking piston 4 as well as the normalized force present at the storagepiston 6 remains constant.

In rotation angle position C, the storage piston 6 has reached the deadcenter of its stroke movement, the end of the filling stroke. At thispoint, the changeover from suction stroke to delivery stroke takes placeas a result of the reversal of movement (lower diagram, FIG. 2). In therange between B and C, the delivery of medium into the high-pressureline 20 is based on the displacement action of the working piston 4only.

Between B and C, the stroke of the working piston 4 in relation to theangle of rotation rises considerably compared to the stroke during thepre-compression phase (O-A) and compared to the phase of common deliveryof the working piston 4 and the storage piston 6 (A-B) (upper diagram,FIG. 2). As can be seen from the lower diagram of FIG. 2, the suctionstroke of the storage piston 6 commences and terminates within thatrange at B and C, respectively. The delivery forces present at theworking piston 4 in the range between B and C correspond to thenormalized forces in the range between A and B, since the working piston4 continues to deliver at system pressure (nominal pressure) on thehigh-pressure side. The lower diagram of FIG. 4 further shows that, dueto the continued exposure to system pressure, the normalized forcetransmitted at the storage piston 6 remains constant also during thesuction stroke. The reason for that being that the storage piston 6 isunder system pressure both on the inlet side and on the outlet side.Referred to a full pumping cycle, the force value curve shows thetheoretical constant course. It runs parallel to the x-axis since thestorage piston is continuously exposed to the full hydrostatic pressuregoverned by the system pressure prevailing beyond of the pumps outlet.

The additional force induced by the friction between the piston and itsseal varies with the system pressure which governs the force pressingthe sealing lip of the seal against the piston surface in the contactarea. Said friction force, however, reaches only a fraction of the forcevalue which is hydrostatically exerted onto the pump structure beingmonitored there. The seal friction force is superimposing thehydrostatic force in additive as well as in subtractive manner. Lookingto actual curves of the force values monitored during a pumping cyclereveals that the superimposition of the friction force manifests in a(very small) stepwise offset whereby the addition or subtraction dependson the direction of the piston stoke movement: An additive effect—shownin FIG. 4—is associated with the displacement stroke (section O-B andC-E) and a subtractive effect is associated with the filling stroke(section B-C).

It can be seen from FIG. 3 that the normalized displacement action ofthe storage piston 6 in the range between B and C (during the suctionstroke) becomes negative. The normalized displacement action of theworking piston 4 exceeds the combined displacement action of bothpistons by this amount, said displacement action being illustrated by adotted line in the diagrams of FIG. 3. In the range between B and C thedisplacement action of the working piston 4 is divided into the amountof medium volume pumped by the storage-piston/cylinder unit 3 into thehigh-pressure line 20 and into the amount of medium volume received inthe storage-piston/cylinder unit 3 during the suction stroke thereof.

In rotation angle position C, the working piston 4 is at its top deadcenter-end of displacement stroke—and the storage piston 6 is at itsbottom dead center—end of the filing stroke. In this rotation angleposition C, the stroke movements of the working piston 4 as well as ofthe storage piston 6 proceed in the respective opposite direction. Thisis apparent, in the upper diagram of FIG. 4, from a steep drop of thenormalized force present at the working piston 4. The delivery phase ofthe working piston 4 terminates upon reaching the angle of rotation C,whereas the displacement action of the storage piston 6 commences,wherein, in the latter phase, the displacement action of the storagepiston in total corresponds to the combined displacement action of theworking piston 4 and the storage piston 6 (FIG. 3). At the angle ofrotation C, the outlet valve 19—associated with the workingpiston—closes and thus prevents a backflow of medium from thehigh-pressure side into the cylinder space 5 of theworking-piston/cylinder unit 2.

The inlet valve 15 opens with delay. This is apparent especially from acomparison of the diagrams of FIG. 6, from which can be gathered thatthe inlet valve 15 opens only at the angle of rotation D. In therotation angle range between C and D, the detrimental dead volume, whichwas not displaced from the working-piston/cylinder unit 2 but beingstill compressed to system pressure during the previous delivery stroke,expands again to the volume at ambient pressure. Consequently, mediumcannot be drawn in from the low-pressure side into the cylinder space 5of the working-piston/cylinder unit 2 during said expansion phase.Accordingly, the normalized force present at the working piston 4substantially reaches a value of 0 only in the rotation angle positionD.

In rotation angle position D, the negative pressure required for theactual initiation of the suction stroke is established in the cylinderspace 5 of the working-piston/cylinder unit 2. In the process the inletvalve 15 opens (upper diagram, FIG. 6) and the suction stroke of theworking piston 4 is performed starting at point D until the top deadcenter is reached at the angle of rotation E. In the range between D andE as well as in the range between C and D the delivery of the mediumfrom the pump into the high-pressure line 20 is based solely on thedisplacement action of the storage piston 6. In rotation angle positionE, which corresponds to the angle of rotation 0, the cycle describedabove is repeated.

A pressure sensor 25 is arranged in the high-pressure line 20 behind thestorage-piston/cylinder unit 3. This sensor is not required forperforming the method according to the present invention and can thus,in principle, also be omitted. In the illustrated embodiment, thepressure sensor 25—as shown by the dotted line—is connected to thecomputing unit 12, preferably via a signalling line. The measuredpressure values provided by the pressure sensor 25, for example, permitconclusions as to the cause of failures and malfunctions, includingthose of the total system 1, by comparing them to measured valuesprovided by the force/moment sensor 13.

In the most preferred embodiment, the invention is carried out on thebasis of a dual piston serial type pump 1 comprising a ‘working piston’4 and ‘a storage piston’ 6, with said pump 1 having a single DC-motor 8paired with a belt drive 40 rotating a dual cam axle 41. The rotationalmotion of the cam disk 42 profiles is converted into linearreciprocating motion by rollers 43 at the near end of Z-shaped drivearms 44. By means of axial ball bearings, the drive arms 44 performtheir reciprocating motions along precision guide axles and at the sametime are stabilised against canting by means of laterally arrangedroller bearings. Furthermore, the drive pistons 4,6, at their distantends, are each fitted with a hard contact rod 45 in a bracketed holderelement 46, with the low friction counter surface of said rod providing,in conjunction with a fork-type spring, a free-floating, no side loadinducing connection between the drive and the pumping piston 4,6.

For allowing continuous monitoring of the liquid displacement process independency of the delivery pressure and the specific compressibility ofthe medium being pumped, the working drive arm 44 is fitted with astrain gauge. By means of the strain gauge 49 the efficiency of thepumping and filling stroke can be continuously and precisely monitoredduring each pumping cycle and, processing its signal track forms thebasis for a fast and efficient feedback control method to compensate forthe impact of the specific compressibility of the pumped medium at highand highest operation pressures on the pumping performance andefficiency, Compare: FIG. 10 a, b.

The pertinent liquid displacement assembly (LDA) is built to sandwichdesign according to U.S. Pat. No. 5,653,876.

The kinematic data of the drive cam disks 42 for the working 4 and thestorage piston 6 are shown in the figures as diagrams of elevationnormalised relative to the maximum stroke amplitude of the workingpiston 4, covering a full pumping cycle) (360°, FIGS. 7 a and 7 b forworking 4 and storage pistons 6 respectively. The pertinent camreference angles A to E are shown for the implemented cam profiles.These figures should be compared with the idealised plots of FIG. 2.

Covering a full pumping cycle)(360°, the actual normalised deliveryrates with the given cam profiles, at constant angular velocity versuscam angular position are shown in FIGS. 8 a and 8 b for working andstorage pistons respectively. The pertinent cam reference angles A to Eare shown for the actual (implemented) cam profiles. These should becompared with the idealised plots of FIG. 3.

Exemplarily, the actually measured angular velocity as a function of thecam angular position is shown in FIG. 9 for the case of a sufficientcompressibility compensation process step during each pumping cycle.Between A and B, i.e. after the appropriate pre-compression level hasbeen reached in A, during the control period, the motor speed is reducedsuch that the cam rotational speed is lowered to approximately 40% and,subsequently, the cam rotational speed is increased again to its initialnominal value in such way that a composite constant delivery flow isgenerated by both pistons.

In an ideal way, deceleration is achieved inherently almostinstantaneously, since the outlet check valve is hydraulically actuatedwith negligible delay, whereas the acceleration to the nominal camvelocity up to point B is synchronized with the cam drive profiles.Feedback trigger events in section A to B are processed in the controlunit as the first derivative of the monitored force at the workingpiston, as a function of the angle of rotation of the cam shaft.

The actual normalised forces, measured with the force sensors, are shownin FIG. 10 once a complete compressibility compensation control cyclehas been achieved by applying the method of control already described insuccessive cycles. Typically, 2 to 3 pumping stroke cycles have to beperformed until steady and continuous flow conditions are established atthe high pressure outlet side of the system.

The shown signal traces are equivalent to the idealised traces in FIG.4. The characteristic negative force of small value from rotation anglesection D to E, measured at the working piston 4, quantifies thefriction force of the piston seal free from hydraulic load, with theinlet check valve 38 opened for the filling stroke. Depending on thesign of the piston movement, the storage piston force sensor measures analternating effective friction due to the contribution caused by thepertinent seal which is permanently under load depending of the systemback pressure. The small deviations of the actual measured valuesrelative to the theoretical values are contributed mainly by non-linearvariations of the seal friction.

As described previously, the seal friction force is superimposing thehydrostatic force in an additive as well as in a subtractive mode.Looking to actual curves of the force values monitored during a pumpingcycle reveals that the superimposition of the friction force manifestsby a comparably small stepwise offset whereby addition or subtractiondepends on the direction of the piston stoke movement: An additiveeffect—shown in FIG. 10 b—is associated with the displacement stroke(section O-B and C-E) and a subtractive effect is associated with thefilling stroke (section B-C).

The actual normalised first derivative of the measured force is shown inFIG. 11. It is equivalent to FIG. 5. The shown raw and un-dampenedsignal trace being noisy, still reveals amplitudes which are well aboverequired levels to discriminate the various phases in the pumping cycle.

In a preferred embodiment a Maxon Motor™ RE 268214 with graphite brushesis used. However any DC motor with adequate power rating, sufficienttorque and high bandwidth rotational speed specifications matching orexceeding the performance values of that motor will be sufficient todrive the pump system described.

The motor 1 is typically operated within the range of a few rpm up to12.000 rpm to generate the pump's flow rate range specified from 0 to 5ml/min maximum.

In FIG. 12 the principal layout of the dual stage gear system isdepicted. Using a commercially available gear, it is built according toknown design. The dynamic velocity and acceleration range required forgenerating the specified flow rate range is met by a dual stage beltdrive system providing a reduction rate of 100:1. By its concept, thegear system corresponds to a conventional and commercially availabledesign.

FIG. 13 shows the gear system including one of the two Z-shaped drivearms 44 or primary pistons in cross-sectional view. The drive pistons 44are fitted with axial ball bearings 47 to perform their reciprocatingmotions along a precision guide axle. They are actuated by the camprofiles 42 via a roller 43 each at their near ends, converting therotating cam motion into reciprocating piston stroke motions. At theirdistant end they bear in a clamping end piece 48 a contact rod 45 fromhard material which exhibits a low friction counter surface for thepumping piston 44. In conjunction with a fork-type spring, said surfaceensures a free-floating connection between the actuation end and thepumping piston shaft 45.

The strain gauge 49 to be seen in the left ‘arm’ of the drive piston 44is fixed with adhesive into a countersunk hole. In scale, the straingauge 49 shown is shown oversized for clarity. In the presentembodiment, the gear system is identical for both the working 4 and thestorage piston 6; hence only one half section of the gear system isshown. Among the possible force sensing devices, the type of forcesensor used in the most preferred embodiment is based on the knownresistive strain gauge principle. The strain gauges 49 are fixed withcured adhesive and they are conveniently located in the cantileveredpart of the Z-shaped drive piston 44, where the flexure and/or shearstrains reach their measurable maximum.

A commercially available and industrial standard resistive strain gaugetype 49 is used, such as the Vishay™ Type 062LV. Said type allowsenables an optimum of shear strain measurement including fully balancedWheatstone bridge arrangements. The force sensor(s) required forimplementing the feedback pump control concept according to theinvention is integrated into the structure of the drive piston 44associated with the working piston 4 and, optionally also in the drivepiston 44 associated with the storage piston 6.

Shown in FIG. 14 is a lengthwise cut in flow direction of the mostpreferred embodiment of the liquid end.

1. A piston pump for generating a substantially pulsation-free deliveryflow of a medium to be delivered, comprising: at least first and secondpiston-cylinder-units for delivering the medium out of a low-pressurearea into a high-pressure area; a drive unit for driving thepiston-cylinder units; a control unit for controlling the driven speedof the drive unit; and a sensor for collecting an actual value of acontrol parameter, by which actual value the extent of pulsation of theflow generated in the high-pressure area is derivable; wherein thesensor is adapted for collecting values of mechanical forces and/ortorques exerted or transmitted in or at the structure of at least onepiston-cylinder unit and/or the drive unit.
 2. The piston pump of claim1, further comprising: an area wetted by the medium being pumped; and adry area, which is not wetted by the medium during operation; whereinthe sensor is located in the dry area of the pump and is not wetted bythe medium to be delivered.
 3. The piston pump of claim 1, wherein themechanical forces include at least one of tensile stress, compressionstress, shear stress and torsional stress.
 4. The piston pump of claim1, wherein the sensor is based on one of a strain gauge, apiezo-electric element, an optical measuring device and an acousticresonator.
 5. The piston pump of claim 1, wherein the piston pump is aparallel dual piston pump.
 6. The piston pump of claim 1, wherein thepiston pump is a serial dual piston pump.
 7. The piston pump of claim 1,wherein each piston-cylinder unit is driven by a separate drive unit. 8.The piston pump of claim 1, wherein each drive unit is driven by aseparate motor.
 9. The piston pump of claim 1, wherein the firstpiston-cylinder unit has a suction side in fluid communication with thelow-pressure area and a discharge side in fluid communication with thesecond piston-cylinder unit and the high-pressure area in parallel. 10.The piston pump of claim 9, wherein a discharge side of the secondpiston-cylinder unit is hydraulically coupled to the high-pressure area.11. The piston pump of claim 9, wherein the sensor monitors mechanicalforces and/or torques being exerted or transferred in or at thestructure or drive of the first piston-cylinder unit.
 12. The pistonpump according to claim 11, further comprising an additional sensorconfigured to monitor mechanical forces and/or torques being exerted ortransmitted in or at the structure or drive of the secondpiston-cylinder unit.
 13. The piston pump of claim 1, wherein the driveunit is designed to generate a pre-compression stroke.
 14. The pistonpump of claim 1, wherein the drive unit comprises a cam drive.
 15. Amethod for controlling a piston pump for delivering a medium out of alow-pressure-area into a high-pressure-area, the piston pump includingat least first and second piston-cylinder units, a drive unit fordriving at least one of the piston-cylinder-units, and a sensor formonitoring mechanical forces and/or torques exerted or transmitted in orat the structure of at least one of the piston-cylinder units and/orsaid drive unit, comprising: i) driving at least one piston-cylinderunit by a drive unit at a first speed (n), and monitoring using thesensor mechanical forces and/or torques exerted or transmitted in or atthe structure of the at least one driven piston-cylinder unit and/or thedrive unit; ii) monitoring the moment of the actual onset of delivery ofthe pumped medium out of the at least one driven piston-cylinder unitinto the high-pressure-area by using the values of the mechanical forcesand/or torques monitored by the sensor; iii) monitoring the rate ofcompression of the pumped medium at the moment of the actual onset ofdelivery; iv) modulating the drive unit rotational speed to a secondspeed (n+1), such that a varying compression rate due to varying systemback pressure and/or varying compressibility of the pumped mediumarising on the high-pressure side of the at least one drivenpiston-cylinder unit is compensated and a substantially pulse-freedelivery flow of the pumped medium is generated in the high-pressurearea and v) repeating steps i) to iv) for each pumping cycle using thedrive unit rotational speed (n+1) modulated in step iv) as the firstspeed (n).
 16. The method of claim 15, wherein in step ii) the variationof the values measured for the mechanical forces and/or torques ismonitored.
 17. The method of claim 15, wherein the mechanical forcesand/or torques or the course of their measuring values in step ii) aremonitored during a single pump stroke at constant drive unit rotationalspeed.
 18. The method of claim 15, wherein stored running programs areused for modulation of the drive unit rotational speed.
 19. The methodof claim 18, wherein a correction factor is derived based on the actualonset of delivery and wherein a specific rotational speed controlprogram is selected depending on the correction factor.
 20. The methodof claim 15, wherein in step i) a rotational speed control program isapplied, which is tuned to the flow rate set at the piston pump.
 21. Themethod of claim 15, wherein the medium being pumped is pre-compressed atthe beginning of each pumping cycle.
 22. The method of claim 21, whereina secondary correction factor is superimposed onto the drive unitrotational speed which compensates for the excess delivery flow referredto volume at ambient pressure generated by the precompression.
 23. Themethod of claim 21, wherein the pre-compression stroke is adapted to aspecified maximum delivery pressure and a maximum specificcompressibility to be expected for the pumped medium.
 24. The method ofclaim 21, wherein the drive unit rotational speed is reduced when beingmodulated during the pre-compression stroke.
 25. The method of claim 15,wherein the composition of medium to be delivered by the pump variesover time.
 26. The method of claim 15 wherein a volume which cannot beexpelled from the displacement area of the piston-cylinder unit of theworking piston as detrimental dead volume, but still has to becompressed according to current operation conditions before actualpumping can onset, is monitored for each pumping cycle in order todetermine the loss of filling efficiency due to expansion of said volumeat the beginning of the suction stroke.
 27. The method of claim 25,wherein the gradient composition for the subsequent filling stroke iscorrected according to the expansion of the not dischargeabledetrimental dead volume within the displacement area of the firstpiston-cylinder unit.
 28. The method of claim 15, wherein the drive unitcomprises a cam drive.